Compressor

ABSTRACT

A dry-compressing compressor comprises two screw rotors in a housing defining a suction chamber. At a compressor inlet of the compressor preferably atmospheric pressure prevails and at a compressor outlet of the compressor preferably a pressure of more than 2 bars (absolute) prevails. For each screw rotor at least one displacement element including a helical recess defining a plurality of windings is provided. The at least one displacement element per screw rotor has a single-pass asymmetrical profile.

BACKGROUND 1. Field of the Disclosure

The disclosure relates to a compressor, in particular a screwcompressor.

2. Discussion of the Background Art

For compressing gases, in particular for providing compressed air,primarily oil-injected screw compressors are nowadays used. They canusually perform a compression from 1 bar (absolute) to 8.5 to 14 bars(absolute) in one compressor stage. Here, the delivered intake volumeflows range from 30 to 5000 m³/h. Such screw compressors comprise twocounter-rotating screw rotors. The screw rotors each comprise at leastone helical deepened portion such that a displacement element is formed.The injection of oil into the suction chamber, where the two screwrotors are arranged, serves for sealing the gaps between the rotors andthe housing and/or the inner wall of the suction chamber. By providingoil, a sufficient tightness can be attained for realizing highcompression pressures of in particular up to 14 bars in one compressorstage. In addition, the oil serves for lubricating the rolling contactsbetween the two screw rotors. Therefore, a synchronization gear for thetwo screw rotors is not required. Further, the oil serves fordischarging compression heat. Only in this manner, a low temperature canbe attained at a high efficiency. Finally, the oil serves for dampingmechanical noise. An essential disadvantage of the use of oil is thatthe oil enters the gas to be delivered. The oil must be removed from thecompressed air with the aid of multi-stage separators. As a result, suchcompressors are complex and require a large installation space. The useof oil-injected screw compressors in particular in areas where a highpurity of the compressed air is required, such as in the field ofpharmaceutical or food industry, is not possible or possible only whenusing extremely complex multi-stage oil separators.

For generating oil-free compressed air, it is known to usedry-compressing screw compressors. Here, the two screw rotors arearranged in a contactless manner and synchronized to each other via anoil-lubricated gear. However, dry-compressing screw compressors have thedrawback that one compressor stage only allows for a compression to 4 to5 bars (absolute). The reason for this is in particular that largeleakages occur through the gaps between the rotors and the housing. Forreaching pressures of 9 bars (absolute), for example, two-stage screwcompressors must therefore be used. Besides the two compressor stages,an intermediate cooling of the compressed air is necessary, whichresults in complex equipment comprising many components and requiring alarge installation space.

In addition, dry-compressing compressors configured as so-called rotarytooth compressors are known. These, too, have the drawback that theymust be of a multi-stage configuration for achieving high pressures ofapproximately 9 bars (absolute).

In addition, dry-compressing spindle compressors are known. Thesecomprise a plurality of closed working chambers arranged one behind theother along a plurality of windings or loops of a displacer.Theoretically, high compression pressures are said to be achieved evenwith a one-stage design such that the spindle compressors can substitutemulti-stage screw compressors or rotary tooth compressors. However,spindle compressors are so far not commercially available such thatthere is no evidence that high compression pressures can be reached witha one-stage design.

Spindle compressors are described in DE 10 2010 064 388, WO 2011/101064,DE 10 2012 202 712 and DE 10 2011 004 960, for example.

It is an object of the disclosure to provide a dry-compressingcompressor with the aid of which high pressures of in particular morethan 5 bars (absolute) can be reached even with a one-stage design.

SUMMARY

The dry-compressing compressor according to the disclosure comprises asuction chamber defined by a housing. In the suction chamber, two screwrotors engaging with each other are arranged. These are counter-rotatedwith respect to each other for delivering the gas. For this purpose,each screw compressor comprises at least one displacement element havinga helical recess for defining the windings. In particular, for eachscrew rotor only one displacement element can be provided which can beintegrally formed with a rotor shaft. Further, the housing comprises acompressor inlet where preferably atmospheric pressure prevails. At acompressor outlet preferably a pressure of more than 2 bars (absolute)prevails, wherein it is particularly preferred that a pressure of morethan 5 bars (absolute) prevails at the compressor outlet.

With the aid of the dry-compressing compressor according to thedisclosure high pressures can be reached with a one-stage design since,according to the disclosure, the at least one displacement element perscrew rotor is of a single-pass configuration and has an asymmetricalprofile. According to a particularly preferred embodiment, theasymmetrical profile is configured such that no or merely a smallblowhole occurs. Since no continuous blowhole exists, in a profile whichis preferably asymmetrical according to the disclosure a short-circuitmerely occurs between two adjacent chambers. According to a particularlypreferred embodiment, the so-called Quimby profile is provided as theasymmetrical profile. Asymmetrical profiles have two different profileedges. Although the manufacture is complex due to the required twoseparate operating steps, an extremely tight working chamber can berealized.

Providing single-pass, possibly even symmetrical rotor profiles offersthe advantage that a larger tightness can be achieved. In the case ofprofiles having two more passes of the respective meshing displacementelements, connections across several chambers are formed through thegaps such that the leakage affects the delivered gas flow and the energyconversion quality.

According to another preferred embodiment of the dry-compressingcompressor according to the disclosure, the number of windings of the atleast one displacement element or, in the case of a plurality ofdisplacement elements the sum of the windings of the displacementelements of a screw rotor is larger than the ratio of the pressureprevailing at the compressor outlet to the pressure prevailing at thecompressor inlet. The number of windings thus results from

$n > \frac{Pout}{Pin}$

wherein p_(out) is the outlet pressure and p_(in) is the inlet pressureof the compressor. It is particularly preferred that the number ofwindings or loops is calculated as follows

$n > {\frac{Pout}{Pin} + 4.}$

Due to such a large number of windings or loops per screw rotor, acontinuous but relatively slow compression of the gas is achieved.Thereby, it is possible to easily discharge heat produced during thecompression.

In addition, it is preferred that the installed volume ratio of thedry-compressing screw compressor between the theoretical delivery volumeat the inlet stage (V_(in)) and the theoretical delivery volume at theoutlet stage (V_(out)) is adapted to the pressure ratios at the inlet(p_(in)) and the outlet (P_(out)). Here, p_(in) and p_(out) are definedas absolute pressures. Preferred is a volume ratio V_(i) of

$V_{i} = {\frac{Vin}{Vout} = \left( \frac{Pout}{Pin} \right)^{1\text{/}k}}$

wherein n has a value of k−0.3 to k+0.3 and preferably a value betweenk−0.1 and k+0.1. Here, k is the isotropic exponent of the gas mixture tobe delivered.

According to another preferred embodiment, the displacement elementscomprise at least one area or portion where the chamber volume Vin de-creases to an intermediate volume V_(VK).

According to another preferred or alternative embodiment, the decreaseof the delivery volume of the stages (working chambers) from the largeinlet volume (V_(in)) to the smaller outlet volume (V_(out)) is dividedinto two areas. Here, it is particularly preferred that in the firstarea the working chamber closed towards the suction side is reduced to aspecific volume (volume of the precompression V_(VK)) within a smallrotation angle range. Here it is preferred that

V _(VK) =x V _(in)

wherein x=0.1 to 0.5, in particular x=0.2 to 0.4, and particularlypreferred x=0.3. Due to the compression operation, the precompressionraises the temperature of the gas to a moderate value of 150° C.-200° C.In the second area of the compression, depending on the rotation angle,the working chamber volume decreases to a considerably smaller extentthan in the first area. The rotation angle and thus the number of stagesin the second area is considerably larger than in the first area. Due tothe moderate temperature rise in the first area, the large housingsurface in the second area and the relatively long dwell time of the gasin the second area due to the larger rotation angle, in the second areaanother temperature rise of the gas due compression can be avoided to alarge extent by heat transport into the housing.

The compression of the gas is selected such that the producedcompression heat can be easily discharged via the side walls of thehousing such that the temperature of the gas does not rise or rises onlyto a small extent. Here, the maximum temperature change is preferablyless than 50° C., and particularly preferably less than 30° C.

A particular advantage of the selected division of the volume decreaseis that a largely homogeneous temperature distribution in the componentis achieved. Thereby, thermal peak loads and the associated largecomponent expansions can be avoided.

The ratio between the inlet volume (V_(in)) and the volume of theprecompression (transition from the first to the second area V_(VK)) canbe related to the internal volume ratio v, of the compressor

$v_{VK} = {\frac{Vin}{Vout} = \left( v_{i} \right)^{1\text{/}j}}$

wherein j=2 to 5, in particular j=2.5 to 3.5, and particularly preferredj=3.

According to a particularly preferred embodiment, the precompression isperformed in the described first area at 1.5 to 3 rotor revolutions(windings).

According to a preferred embodiment, the inventive large number ofwindings in the second area can be realized by a single displacementelement for each rotor. However, it is also possible to provide acorresponding number of windings in this discharge-side area by twodisplacement elements, for example. By providing an inventive largenumber of windings in this area, where, according to the disclosure,preferably the medium to be delivered is only compressed to a smallextent per winding, it is possible to do without internal cooling of therotors. The reason for this is in particular that due to the relativelysmall extent of compression in this area the temperature increase of thedisplacement element caused by compression is small. In addition, inthis area, due to the high density of the delivered medium, a good heatdissipation from the displacement element into the compressor housingvia the medium is realized.

Preferably, the screw rotors and the at least one provided displacementelement are configured such that between an area where 5%-20% of theoutlet pressure prevails, and the discharge-side rotor end at least 6,in particular at least 8, and particularly preferably at least 10windings are provided. Here, the discharge-side rotor end is the area ofthe compressor outlet. Here, according to a preferred embodiment, theinventive large number of windings in this area can be provided at asingle discharge-side displacement element provided per rotor. However,it is also possible to provide a corresponding number of windings inthis discharge-side area at two displacement elements, for example. Byproviding an inventive large number of windings in an area where,according to the disclosure, the medium to be delivered is onlycompressed to a relatively small extent, it is possible to do withoutinternal cooling of the rotors. The reason for this is in particularthat due to the relatively small extent of compression in this area thetemperature increase of the displacement element caused by compressionis smaller. In addition, in this area, due to the high density of thedelivered medium, good heat dissipation from the displacement elementinto the compressor housing via the medium is realized.

Moreover, due to the preferably large number of winding, a large surfacearea for heat exchange to the housing is available.

It is particularly preferred that the preferably at least 6, inparticular at least 8, and particularly preferably at least 10 windingsare provided in a discharge-side displacement element.

In addition, for configuring screw rotors without internal coolingaccording to the disclosure, it is preferred that the discharge-sidedisplacement element has a mean working pressure of more than 2 bars(absolute) at at least 6, in particular at least 8, and particularlypreferably at least 10 windings. In particular, it is intended torealize a flat pressure gradient inside the compressor. Therefore, thepressure should slowly rise across many windings, in particular 6 to 10windings.

According to the disclosure, it is thus preferably possible to provide acold gap having a height of 0.03 mm-0.2 mm, and in particular 0.05mm-0.1 mm between the surface of the at least one displacement elementand the inside of the section chamber, in particular in thedischarge-side area, even in the case of rotors without internal coolingof the rotors or a housing of aluminum or an aluminum alloy. Such arelatively large gap height can be provided due to the inventiveconfiguration of the particularly 6, preferably 8, and particularlypreferably 10 last windings, as described above.

According to another preferred embodiment of the disclosure, a relativelong screw rotor relative to the diameter is selected. In particular,the at least one displacement element per screw rotor or, in the case ofa plurality of displacement elements per screw rotor, said plurality ofdisplacement elements jointly have a ratio of length L to diameter Dwhere the following applies:

$\frac{L}{D} > {\frac{Pout}{2{Pin}} - 2}$ and  in  particular$\frac{L}{D} > {\frac{Pout}{2{Pin}} - 1}$

By providing a long rotor having in particular many chambers, the areausable for heat dissipation is increased. Due to the resulting good heatexchange, the gas temperatures of the compressed gas are relatively low.Providing many chambers further offers the advantage that the pressuredifferences between adjacent chambers are small and thus a largetightness can be achieved. Due to such a reduction of the deliveryvolume per stage from the inlet to the outlet side, the compressionprocess becomes particularly effective in terms of thermodynamics andthe gas temperatures remain relatively low. Here, it is particularlypreferred that the internal volume ratio is adapted to the ratio ofoutlet to inlet pressure such that neither overcompression orcompression by re-aeration occurs.

The internal volume ratio can be attained by varying the pitch of thewindings. Preferably, the pitch of the windings is in particular changedsuch that it is decreased and/or becomes steeper from the compressorinlet to the compressor outlet. The pitch can be changed continuouslyand/or stepwise.

In addition to or instead of the variation of the pitch, the head orfoot diameter of the profile can be changed continuously or stepwise.Again, a continuous change of the head or foot diameter is particularlypreferred such that the rotor has a conical configuration, in particularin combination with a continuous change of the pitch.

According to a particularly preferred embodiment, the pressure ratiobetween the outlet pressure and the inlet pressure is at least 5.According to a particularly preferred embodiment, the outlet pressure isat least 2 bars (absolute), in particular at least 5 bars.

According to another particularly preferred embodiment, thedry-compressing compressor comprises at the compressor inlet and/or atthe compressor outlet a respective gas collection chamber preferablyinside the compressor housing.

Moreover, it is preferred that the dry-compressing compressor is acompressor having two shafts. The latter are preferably supported onboth sides such that narrow gaps can be realized both between thedisplacement elements and between the displacement elements and theinner wall of the suction chamber. Preferably, the two rotor shafts aresynchronized by a synchronization gear preferably arranged outside thesuction chamber. The bearings can be lubricated by grease and/or oil.Likewise, the gear can be lubricated by grease and/or oil. This ispossible since both the bearings and the synchronization gear arepreferably arranged outside the suction chamber and it is thus avoidedthat the gas to be delivered is contaminated by oil.

Preferably, the housing is made from aluminum or an aluminum alloy.Here, an aluminum alloy AlSi7Mg or AlMg07,5Si for the housing isparticularly preferred. In particular, the heat expansion coefficient(expansion coefficient) of the material of the screw rotors is smallerthan the expansion coefficient of the material of the housing. It isparticularly preferred that the expansion coefficient of the screwrotors is smaller than 12*10⁻⁶1/K. This can be achieved with rotors madefrom iron or steel materials.

The two screw rotors arranged in the suction chamber comprise at leastone displacement element having a helical recess. The helical recessesdefine several windings. According to the disclosure, the at least onedisplacement element is made from a steel or iron alloy. It is thusparticularly preferred that the screw rotors including the displacementelements are made from a steel or iron alloy. The housing is also madefrom a steel or iron alloy or from aluminum or an aluminum alloy.

Preferably, each displacement element comprises at least one helicalrecess having the same contour along its overall length. Preferably, thecontours are different for each displacement element. Thus theindividual displacement element preferably has a constant pitch and anunvarying contour. Thereby, the manufacture is considerably simplifiedsuch that the manufacturing costs can be largely reduced.

For further improving the suction capacity, the contour of thesuction-side displacement element, that is in particular the firstdisplacement element as seen in the pumping direction, is preferably ofan asymmetrical configuration. Due to the asymmetrical configuration ofthe contour and/or the profile, the edges can be configured such thatthe leakage areas, the so-called blowholes, can in particular completelydisappear or have at least a smaller cross-section. A particularlysuitable asymmetrical profile is the so-called “Quimby” profile.Although such a profile is relatively difficult to produce, it offersthe advantage that no continuous blowhole exists. A short-circuit occursonly between two adjacent chambers. Since this is an asymmetricalprofile having different profile edges, at least to working steps arerequired for the production since the two edges have to be produced intwo different working steps due to their asymmetry.

The discharge-side displacement element, in particular the lastdisplacement element as seen in the pumping direction, preferably has asymmetrical contour. The symmetrical contour in particular offers theadvantage that it is easier to produce. In particular, the two edgeshaving a symmetrical contour can be produced in one working step using arotating end milling cutter or a rotating side milling cutter. Althoughsuch symmetrical profiles have blowholes, these are continuous, i.e. donot only exist between two adjacent chambers. The size of the blowholedecreases with decreasing pitch. Thus, such symmetrical profiles can inparticular be provided for the discharge-side displacement elementsince, according to a preferred embodiment, it has a smaller pitch thanthe suction-side displacement element and preferably also than thedisplacement element arranged between the suction-side and thedischarge-side displacement element. Even though the tightness of suchsymmetrical profiles is somewhat smaller, they offer the advantage thatthey are considerably easier to produce. In particular, it is possibleto produce the symmetrical profile in a single working step andpreferably using a simple end milling cutter or side milling cutter.Thereby, the costs are considerably reduced. A particularly suitablesymmetrical profile is the so-called “cycloidal profile”.

Providing at least two such displacement elements results in thecorresponding screw compressor being capable of generating high outletpressures at a low power consumption. Further, the thermal load issmall. Arranging at least two displacement elements having theconfiguration according to the disclosure with a constant pitch and anunvarying contour in a compressor leads to essentially the same resultsas with a compressor having a displacement element with a varying pitch.At high installed volume ratios three or four displacement elements perrotor can be provided.

According to a particularly preferred embodiment, for increasing theattainable outlet pressure and/or for reducing the power consumptionand/or the thermal load, a discharge-side displacement element, that isin particular the last displacement element as seen in the pumpingdirection, comprises a large number of windings. A large number ofwindings allows for accepting a larger gap between the screw rotor andthe housing at constant performance. Here, the gap can have a cold-gapwidth of 0.05-0.3 mm. A large number of outlet windings or of windingsof the discharge-side displacement element is inexpensive to producesince, according to the disclosure, this displacement element can have aconstant pitch and preferably also a symmetrical contour. On the outletside an asymmetrical profile can be used. This allows for an easy andinexpensive production such that it is acceptable to provide a largernumber or windings. Preferably, this discharge-side or last displacementelement has more than 6, in particular more than 8, and particularlypreferably more than 10 windings. According to a particularly preferredembodiment, the use of symmetrical profiles offers the advantage thatboth edges of the profile can be simultaneously cut with a millingcutter. Here, the milling cutter is supported by the respective oppositeedge such that deformation or distortion of the milling cutter duringthe milling operation and resultant inaccuracies are avoided.

For further reducing the manufacturing costs it is particularlypreferred to integrally form the displacement elements and the rotorshaft.

According to another preferred embodiment, the change of pitch betweenadjacent displacement elements is inconsistent or erratic. Possibly, thetwo displacement elements are arranged at a distance to each other inthe longitudinal direction such that between two displacement elements acircular cylindrical chamber is defined which serves as a tool outlet.This is in particular advantageous for manufacturing integrally formedrotors since the tool producing the helical line can be easily removedin this area. If the displacement elements are manufactured separatelyfrom each other and are then mounted to a shaft, a tool outlet, inparticular such an annularly cylindrical area need not be provided.

According to a preferred aspect of the disclosure, no tool outlet isprovided between two adjacent displacement elements at the locationwhere the pitch changes. In the area of the change of pitch both edgespreferably have a discontinuity or recess for removing the tool. Such adiscontinuity has no notable influence on the compression capacity ofthe compressor since it is a localized discontinuity or recess.

The compressor screw rotor according to the disclosure in particularcomprises a plurality of displacement elements. These may have the sameor a different diameter. Here, it is preferred that the discharge-sidedisplacement element has a smaller diameter than the suction-sidedisplacement element.

In the case of displacement elements manufactured separately from therotor shaft, the former are mounted to the shaft by press-fitting. Here,it is preferred to provide elements, such as locating pins, for definingthe angular position of the displacement elements relative to eachother.

It is particularly preferred that the screw rotor is formed integrallyin particular from a steel or an iron alloy. The screw rotor can alsocomprise a rotor shaft which supports the at least one displacementelement. In particular when providing a plurality of displacementelements, this offers the advantage that these can be manufacturedseparately from each other and then be connected to the rotor shaft inparticular by press-fitting or shrink-fitting. Here, it is possible toprovide fitting keys or the like for defining the angular position ofthe individual displacement elements.

If a plurality of displacement elements per screw rotor are provided, itis possible to integrally form the displacement elements.

According to the disclosure, it is preferred that the screw rotors haveno internal cooling. Hence it is particularly preferred that the screwrotors do not have any ducts through which in particular liquid coolantflows. However, the screw rotors can comprise boreholes or ducts for thepurpose of weight reduction, for balancing or the like, for example. Itis particularly preferred that the screw rotors are of a solidconfiguration.

In addition, it is preferred that the housing has a mean heat flowdensity in the area of the displacement elements of less than 80,000W/m², preferably less than 60,000 W/m², and in particular less than40,000 W/m².

The mean heat flow density is the ratio of the compression capacity tothe wall surface of the compression area.

In the dry-compressing screw compressor according to the disclosure agas aftercooler and/or a condensate separator for separating thecondensate produced by compression and/or a silencer may additionally beprovided at the compressor outlet. Further, it is possible to provide aninlet air filter or an inlet silencer at the compressor inlet.

Particularly preferably, with the aid of the compressor according to thedisclosure a volumetric efficiency of at least 70 percent, preferably atleast 85 percent for at least one operating point of the compressor canbe achieved. A decisive factor is the ratio of theoretically possibleand practically achieved volume flow. The high volumetric efficiencyadapted to be achieved by the compressor according to the disclosure isan indication of the good tightness of the compressor.

Further, the compressor according to the disclosure preferably has ahigh isothermal efficiency factor of at least 45 percent, preferably atleast 60 percent. The isothermal efficiency factor is the ratio of idealisothermal compression capacity and real compression capacity. Theisothermal efficiency factor is also an indication of good tightness andgood cooling of the compressor.

In addition, it is preferred that the dry-compressing compressor isoperated by a motor at a mean speed. In particular, the speed is higherthan 3,000 1/min, and particularly preferred more than 4,000 1/min. Onthe other hand, the speed is preferably lower than 10,000 1/min.

At relatively low speeds in the range of 3,000 1/min of conventionalasynchronous motors, for example, large rotor diameters must be used.This results in unfavorable ratios of delivered gas volume and leakageareas. This is approximately proportional to the rotor diameter. On theother hand, very high speeds of more than 10,0000 1/min entail very highdemands on the balancing of the rotors or the displacement elements.This is difficult to achieve in the case of single-pass screw threads.In addition, with increasing power density due to high speeds, itbecomes more and more difficult to cool the compressor. Another drawbackof very high speeds with very small tooth gaps is the high gas frictionin the gas paths. Thereby, the energy efficiency decreases. At meanspeeds according to the disclosure a good compromise between tightness,balancing, gas friction and heat transfer or temperature level can beachieved.

Preferably, the housing is intensively cooled for keeping the gas andthe components cool. In the embodiment of the compressor according tothe disclosure, this can possibly also be achieved without internalcooling of the rotors. Low gas temperatures cause a reduction of thecompression operation and thus have a positive effect on the powerconsumption of the compressor.

According to a preferred aspect of the disclosure, the rotors and/or thedisplacement elements can be coated with layers on the basis of PTFE ormolybdenum sulfide, for example, in order to decrease the gap heightswithout affecting the operational safety.

BRIEF DESCRIPTION OF THE DRAWINGS

Hereunder the disclosure will be explained in detail on the basis of apreferred embodiment with reference to the accompanying drawings inwhich:

FIG. 1 shows a schematic top view of a preferred embodiment of a screwrotor of the screw compressor according to the disclosure,

FIG. 2 shows a schematic sectional view of displacement elements havingan asymmetrical profile,

FIG. 3 shows a schematic sectional view of displacement elements havinga symmetrical profile, and

FIG. 4 shows a schematic sectional view of a screw compressor.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

The screw rotors illustrated in FIGS. 1 to 3 can be used in a screwcompressor according to the disclosure as shown in FIG. 4.

According to a preferred embodiment of the screw compressor, the rotorhas a pitch changing and/or variable in the direction of compression,i.e. from left to right in FIG. 1. In a first suction-side area 10defining a first displacement element a large pitch of approximately50-150 mm/revolution is provided. Here, the pitch changes in the area10, i.e. in the precompression area, to 55-65% of the inlet pitch, i.e.approximately 30-100 mm/revolution. In a second discharge-side area 12corresponding to a second displacement element 12 the pitch isconsiderably smaller. In this area the pitch is in the range of 10-30mm/revolution. In the illustrated embodiment, the at least onedisplacement element per screw rotor is thus defined by a screw rotorhaving a variable, preferably continuously changing pitch. Thiscorresponds to a plurality of displacement elements arranged one behindthe other as seen in the direction of delivery.

In the illustrated preferred embodiment, both in the inlet area and theoutlet area a gas collection chamber 14 each is provided.

Further, the integral screw rotor comprises two bearing seats 16 and ashaft end 18. The shaft end 18 has connected thereto a gearwheel fordriving purposes, for example.

Likewise, it is possible that the individual displacement elements 10,12 are manufactured separately from each other and are separatelyaffixed to the rotor shaft by pressing, for example. Here, the bearingseats 16 and the shaft ends 18 can be integral components of the shaft20. Here, the continuous shaft 20 can be made from a material differingfrom that of the displacement elements 10, 12.

In addition, conical rotors can be provided. According to thedisclosure, they comprise a plurality of displacement elements. Here,too, it is particularly preferred that the plurality of displacementelements are realized by a variable pitch. Conical rotors, too, are of asingle-pass configuration.

FIG. 2 shows a schematic sectional view of an asymmetrical profile (e.g.a Qumiby profile). The illustrated asymmetrical profile is a so-calledQuimby profile. The sectional view shows two screw rotors which meshwith each other and whose longitudinal direction is perpendicular to thedrawing plane. The counter-rotation of the rotors is indicated by twoarrows 15. Relating to a plane 17 extending perpendicularly to thelongitudinal axis of the displacement elements, the profiles of theedges 19 and 21 are of different configuration for each rotor. Theopposing edges 19, 21 must thus be manufactured separately from eachother. However, this somewhat more complex and difficult manufactureoffers the advantage that no continuous blowhole exists but ashort-circuit occurs merely between two adjacent chambers.

Preferably, such an asymmetrical profile is provided for thesuction-side displacement element 10.

The schematic sectional view in FIG. 3 shows a cross-section of twodisplacement elements and/or two screw rotors which are againcounter-rotating (arrows 15). Relating to the symmetry axis 17, theedges 23 of each displacement element are of a symmetricalconfiguration. The preferred exemplary embodiment of a symmetricalcontour illustrated in FIG. 4 is a cycloid profile.

A symmetrical profile, as illustrated in FIG. 3, is preferably providedfor the discharge-side displacement elements 12.

Further, it is possible that more than two displacement elements areprovided. They can possibly have different head diameters andcorresponding foot diameters. Here, it is preferred that a displacementelement having a larger head diameter is arranged at the inlet, i.e. onthe suction side, for realizing a larger suction capacity in this areaand/or increasing the installed volume ratio. Further, combinations ofthe embodiments described above are possible. For example, one or aplurality of displacement elements can be integrally formed with theshaft, or an additional displacement element can be separatelymanufactured and then mounted to the shaft.

In the schematic view of a preferred embodiment of a screw compressoraccording to the disclosure illustrated in FIG. 4, two screw rotors, asillustrated in FIG. 1, are arranged in a housing 26. The compressorhousing 26 comprises an inlet 28 through which gas is taken in in thedirection indicated by an arrow 30. Further, the compressor housing 26comprises a discharge-side outlet 32 through which the gas is dischargedin the direction indicated by an arrow 38. Preferably, the screwcompressor according to the disclosure compresses air in a compressedair chamber.

Between upper surfaces 42 of the two displacement elements 12 and aninner surface 44 of a suction chamber 46 defined by the compressorhousing 26, a gap is formed whose height preferably lies in the range of0.03 mm-0.2 mm and in particular in the range from 0.05 mm-0.1 mm.

The gap between the edges of the displacement elements preferably has agap height of 0.1-0.3 mm.

In the illustrated exemplary embodiment, the compressor housing 26 isclosed by two housing covers 47. The left housing cover 47 in FIG. 4comprises two bearing supports where a ball bearing 48 each forsupporting the two rotor shafts is arranged. On the right side in FIG.4, journals 50 of the two screw rotor shafts protrude through the covers47. On the outside a respective gearwheel 52 is arranged on the twoshaft journals 50. In the illustrated exemplary embodiment, the twogearwheels 52 mesh with each other for synchronizing the two screwrotors with each other. Further, in the right cover 47 in FIG. 4, twobearings 48 for supporting the screw rotors are arranged. In the housingwalls 47 a seal not illustrated is provided in addition to the bearings48.

The lower shaft in FIG. 4 is a drive shaft connected to a drive motornot illustrated.

What is claimed is:
 1. A dry-compressing compressor comprising a housingdefining a suction chamber and having a compressor inlet wherepreferably atmospheric pressure prevails and a compressor outlet wherepreferably a pressure of at least 2 bars (absolute), preferably at least5 bars (absolute) prevails, two screw rotors arranged in the suctionchamber and each having at least one displacement element including ahelical recess for defining a plurality of windings, wherein at leastone displacement element per screw rotor has a single-pass asymmetricalprofile, the screw rotors have no internal cooling of the rotors, andthe housing has a mean heat flow density of less than 80000 W/m² in thearea of the displacement elements.
 2. The dry-compressing compressoraccording to claim 1, wherein the profiles are configured such that notblowhole is formed.
 3. The dry-compressing compressor according to claim1, wherein the profiles of the at least one displacement element of eachscrew rotor are configured a Quimby profile.
 4. The dry-compressingcompressor according to claim 1, wherein a displacement element arrangednear the outlet of the vacuum pump has symmetrical profile.
 5. Thedry-compressing compressor according to claim 1, wherein at least onedisplacement element per screw rotor and/or in the case of a pluralityof displacement elements per screw rotor said displacement elementsjointly comprise a number (n) of windings which is larger than the ratioof outlet pressure (p_(out)) to inlet pressure (p_(in)) such that$n > \frac{Pout}{Pin}$ preferably $n > {\frac{Pout}{Pin} + 4.}$ applies.6. The dry-compressing compressor according to claim 1, wherein theinstalled volume ratio between the delivery volume of the inlet stage(V_(in)) and the outlet stage (V_(out)) is adapted to the pressure ratiobetween inlet pressure (p_(in)) and outlet pressure (p_(out)) such thatthe following applies:$V_{i} = {\frac{Vin}{Vout} = \left( \frac{Pout}{Pin} \right)^{1\text{/}k}}$wherein n has a value of k−0.3 to k+0.3 and k is the isotropic exponentof the gas mixture to be delivered.
 7. The dry-compressing compressoraccording to claim 1, wherein the displacement elements include at leastone area where the volume of the inlet stage (V_(in)) decreases to aprecompression volume (V_(VK)) in a small rotation angle range, whereinthe ratio between inlet volume (V_(in)) and the volume of theprecompression (V_(VK)) is related to the internal volume ratio (v_(i))of the compressor$v_{VK} = {\frac{Vin}{Vout} = \left( v_{i} \right)^{1\text{/}j}}$wherein j=2 to
 5. 8. The dry-compressing compressor according to claim7, wherein the compression from the inlet volume (V_(in)) to theprecompression volume (V_(VK)) takes place during one and a half tothree rotor revolutions (windings).
 9. The dry-compressing compressoraccording to claim 1, wherein at least one displacement element perscrew rotor and/or in the case of a plurality of displacement elementsper screw rotor said displacement elements jointly have a ratio oflength (L) to diameter (D) for which the following applies$\frac{L}{D} > {\frac{Pout}{2{Pin}} - 2}$ and  in  particular$\frac{L}{D} > {\frac{Pout}{2{Pin}} - 1}$
 10. The dry-compressingcompressor according to claim 1, wherein the pitch of the windings ofthe displacement elements varies, preferably changes and particularlypreferably decreases from the compressor inlet to the compressor outlet.11. The dry-compressing compressor according to claim 1, wherein thehead and the foot diameter of the rotor preferably continuously changes,wherein the rotor is in particular of a conical configuration.
 12. Thedry-compressing compressor according to claim 1, wherein the pressureratio $\frac{Pout}{Pin}$ between outlet and inlet pressure is at least5.
 13. The dry-compressing compressor according to claim 1, wherein twoscrew rotors with parallel axes are provided.
 14. The dry-compressingcompressor according to claim 1, wherein at the compressor inlet inparticular inside the housing a gas collection chamber is provided. 15.The dry-compressing compressor according to claim 1, wherein at thecompressor outlet a gas collection chamber is provided in particularinside the housing.
 16. The dry-compressing compressor according toclaim 1, wherein in the housing roller bearings and preferably seals arearranged on both sides of the two screw rotors.
 17. The dry-compressingcompressor according to claim 1, wherein for synchronizing the two screwrotors a synchronization gear is provided.
 18. The dry-compressingcompressor according to claim 1, wherein the speed of the screw rotorsis higher than ${3\text{,}000\mspace{14mu} \frac{1}{\min}},$$\frac{1}{\min},$ $\frac{1}{\min}.$
 19. The dry-compressing compressoraccording to claim 1, wherein the one displacement element is configuredas a discharge-side displacement element and for each screw rotor atleast one further displacement element is provided.
 20. Thedry-compressing compressor according to claim 1, wherein between anupper surface of the displacement element and an inner surface of thesuction chamber a gap with a height of 0.03 mm to 0.2 mm is formed. 21.The dry-compressing compressor according to claim 1, wherein thesuction-side displacement elements have a constant pitch along theiroverall length.
 22. The dry-compressing compressor according to claim 1,wherein each screw rotor comprises a rotor shaft supporting the at leastone displacement element.
 23. The dry-compressing compressor accordingto claim 1, wherein the displacement elements of a screw rotor are of anintegral configuration.
 24. The dry-compressing compressor according toclaim 1, wherein the screw rotors and in particular the at least onedisplacement element per screw rotor have a smaller expansioncoefficient that the housing, wherein the expansion coefficient of thehousing is in particular at least larger than that of the screw rotorsand/or the at least one displacement element.
 25. The dry-compressingcompressor according to claim 1, wherein the screw rotors do notcomprise any ducts through which in particular a liquid coolant flows.26. The dry-compressing compressor according to claim 1, wherein thescrew rotors are of a solid configuration.
 27. The dry-compressingcompressor according to claim 1, wherein a temperature difference in thearea of the discharge-side displacement elements between the latter andthe housing during normal operation is smaller than 50 K.
 28. Thedry-compressing compressor according to claim 1, wherein the distancebetween the area where 5 ° A) to 20% of the outlet pressure prevails andthe last winding of the discharge-side displacement element is at least20% to 30% of the rotor length.
 29. The dry-compressing compressoraccording to claim 1, wherein a gap between the edges of at least one ofthe displacement elements preferably has a gap height of 0.1 to 0.3 mm.